System and method for generating power using a supercritical fluid

ABSTRACT

A dual cycle system for generating shaft power using a supercritical fluid and a fossil fuel. The first cycle is an open, air breathing Brayton cycle. The second cycle is a closed, supercritical fluid Brayton cycle. After compression of air in the first cycle, the compressed air flows through a first cross cycle heat exchanger through which the supercritical fluid from the second cycle flows after it has been compressed and then expanded in a turbine. In the first cross cycle heat exchanger, the compressed air is heated and the expanded supercritical fluid is cooled. Prior to expansion in a turbine, the compressed supercritical fluid flows through a second cross cycle heat exchanger through which also flows combustion gas, produced by burning a fossil fuel in the compressed air in the first cycle. In the second cross cycle heat exchanger, the combustion gas is cooled and the compressed supercritical fluid is heated.

RELATED APPLICATIONS

This application is a divisional application of U.S. patent applicationSer. No. 13/679,856, filed Nov. 16, 2012, now allowed, entitled SystemAnd Method For Generating Power Using A Supercritical Fluid, that claimsbenefit and priority to U.S. Provisional Application Ser. No.61/632,030, filed Jan. 17, 2012; U.S. Provisional Application Ser. No.61/686,043, filed Mar. 29, 2012; U.S. Provisional Application Ser. No.61/688,310, filed May 11, 2012; and U.S. Provisional Application Ser.No. 61/741,303, filed Jul. 17, 2012, the disclosure of each of which ishereby incorporated by reference in its entirety.

TECHNICAL FIELD

The present invention relates to systems and methods for generatingshaft power, especially systems and methods using fossil fuel and aclosed, supercritical fluid cycle.

BACKGROUND

Traditionally, thermodynamic power generation cycles, such as theBrayton cycle, employ an ideal gas, such as atmospheric air. Such cyclesare open in the sense that after the air flows through the components ofthe cycle, it is exhausted back to atmosphere at a relatively hightemperature so that a considerable amount heat generated by thecombustion of fuel is lost from the cycle. A common approach tocapturing and utilizing waste heat in a Brayton cycle is to use arecuperator to extract heat from the turbine exhaust gas and transferit, via a heat exchanger, to the air discharging from the compressor.Since such heat transfer raises the temperature of the air entering thecombustor, less fuel is required to achieve the desired turbine inlettemperature. The result is improved thermal efficiencies for the overallthermodynamic cycle and generally results in efficiencies as high asabout 40%. Larger turbines with more advanced blade aerodynamic designmay achieve even greater efficiencies. However, even in such recuperatedcycles, the thermal efficiency is limited by the fact that the turbineexhaust gas temperature can never be cooled below that of the compressordischarge air, since heat can only flow from a high temperature sourceto a low temperature sink. This is exacerbated by the fact thatemploying higher pressure ratios, which improves the efficiency of theturbine overall, results in higher compressor discharge temperature and,therefore, less heat recovery in the recuperator.

More recently, interest has arisen concerning the use of supercriticalfluids, such as supercritical carbon dioxide, in closed thermodynamicpower generation cycles. Advantageously, supercritical fluids—that is, afluid at or above the “critical point” at which the liquid and gaseousphases are in equilibrium—have a density and compressibility approachingthat of a liquid so that the work required to compress the fluid to thedesired pressure ratio is much lower than it would be for an ideal gas,such as air.

Unfortunately, supercritical fluid cycles suffer from severaldisadvantages that have limited their use. First, although supercriticalfluid cycles are generally closed in the sense that the supercriticalfluid is returned to the cycle inlet after generating power, all of theheat necessary to return the supercritical fluid to near its criticalpoint prior to reintroduction into the compressor cannot be efficientlyconverted to power, so that the supercritical fluid must be cooled bythe transfer of heat to an external heat sink, prior to itsreintroduction into the compressor. This cooling results in the loss ofheat from the cycle and a degradation in thermal efficiency.

Second, unlike what is typically done in air-based open cycles, fossilfuel cannot be combusted in a supercritical fluid without the additionof an oxidizer and subsequent removal of the by-products of combustionfrom the closed cycle. Consequently, supercritical fluids have mostoften been proposed for use in conjunction with nuclear power plants inwhich the nuclear reaction provides the source of heat. Although it ispossible to heat the supercritical fluid in a heat exchanger suppliedwith combustion gas from a conventional fossil fuel fired gas turbine,because of the inefficiency discussed above associated with the highrecuperated compressor discharge temperature and the limited ability totransfer heat into the cycle from the combustion products, the use ofrelatively expensive fossil fuel to heat the supercritical fluid makesthe use of such fuels impractical.

Third, the high pressure of the supercritical fluid, e.g., over 7.0 MPa,creates difficulties in sealing the shafting that transmits the torquedeveloped by the supercritical fluid turbine. If the supercritical fluidcycle is used to generate electrical power, one approach is to includethe electrical generator in the pressure vessel along with the turbineso that the power shaft need not penetrate the pressure vessel. However,this approach has a number of drawbacks. For example, it results in highwindage losses in the generator and requires oil-less bearings.Moreover, maintenance and servicing of the electrical generator becomesmore difficult. Additionally, large generators would require largepressure vessels for containment, resulting in substantial costs andcreating additional points of failure. Also, such an approach cannot beused in applications in which the goal is not the production ofelectrical power, such as in any kind of vehicle propulsion (i.e.turboprop/turbofan applications, automotive and long haul truck drives,marine propulsion) and other applications like oil and gas industryapplications including gas line booster compressors.

Fourth, the efficiency of a supercritical fluid cycle is greatlyaffected by slight deviations in the temperature of the supercriticalfluid in the vicinity of the critical temperature. However, it isdifficult to measure the temperature of the fluid with the requisiteaccuracy to ensure operation at maximum efficiency.

Finally, prior art supercritical carbon dioxide Brayton cycles typicallymake use of recuperation as described above; the reason being thatturbine exhaust temperatures in SCO2 cycles are still very elevated andcompressor discharge temperatures very low making for an ideal recipefor recuperation. This is another reason that SCO2 Brayton cycles are soefficient in nuclear and solar applications. Unfortunately, if a fossilfuel were used as the heat source, passing recuperated compressordischarge through a heat exchanger would make it difficult to pass heatinto the SCO2 flow because the incoming temperature is already so high.

Therefore, the need exists for a system and method for efficiently usinga supercritical fluid in a thermodynamic cycle operating on a fossilfuel and generating shaft power and/or hot water. The need also existsfor an apparatus and method for effectively transmitting torque from theshaft of a supercritical fluid turbine. Further, the need exists for anaccurate method of measuring the temperature of the supercritical fluidin the vicinity of the critical point.

SUMMARY

The present invention encompasses a method of generating shaft power ina system comprising an air cycle and supercritical fluid cycle. Themethod includes the steps of (a) burning a fossil fuel in air so as toproduce a combustion gas, (b) expanding the combustion gas in at least afirst turbine so as to produce an expanded combustion gas, with theexpansion of the combustion gas generating shaft power, (c) compressinga supercritical fluid in a first compressor, (d) flowing at least aportion of the compressed supercritical fluid and the combustion gasthrough the first cross cycle heat exchanger so as to transfer heat fromthe combustion gas to the compressed supercritical fluid so as toproduce a heated compressed supercritical fluid, (e) expanding at leasta portion of the heated compressed supercritical fluid in a secondturbine so as to produce an expanded supercritical fluid, with theexpansion of the supercritical fluid generating additional shaft power,and (f) flowing at least a portion of the expanded supercritical fluidand the air through the second cross cycle heat exchanger prior toburning the fossil fuel in the air so as to transfer heat from theexpanded supercritical fluid to the air. According to one embodiment ofthe invention, the method further comprises compressing the air in asecond compressor so as to produce compressed air prior to burning thefossil fuel in the air so that the fossil fuel is burned in thecompressed air and in which the compressed air flows through the secondcross cycle heat exchanger so as to transfer heat from the expandedsupercritical fluid to the compressed air.

The invention also encompasses a method for generating shaft power in asystem comprising a supercritical fluid cycle and an air cycle thatcomprises the steps of (a) burning a fossil fuel in air so as to producea combustion gas, (b) compressing a supercritical fluid in a firstcompressor, (c) transferring heat from the combustion gas to thecompressed supercritical fluid so as to produce a cooled combustion gasand a heated compressed supercritical fluid, (d) expanding at least aportion of the heated compressed supercritical fluid in a first turbineso as to produce an expanded supercritical fluid, with the expansion ofthe supercritical fluid generating shaft power, (e) returning theexpanded supercritical fluid to the first compressor, and (0transferring heat from the expanded supercritical fluid to the air so asto cool the supercritical fluid to approximately its criticaltemperature prior to burning the fossil fuel in the air and prior toreturning the supercritical fluid to the first compressor. In oneembodiment of the invention, the method further comprises transferringheat from the cooled combustion gas to a flow of water so as to producea flow of heated water.

The invention also encompasses a method for generating shaft power in asystem comprising two supercritical fluid cycles and an air cycle thatcomprises the steps of (a) burning a fossil fuel in air so as to producea combustion gas, (b) compressing a first flow of supercritical fluid ina first compressor so as to produce a first flow of compressedsupercritical fluid, (c) transferring heat from the combustion gas tothe first flow of the compressed supercritical fluid so as to produce acooled combustion gas and a first flow of heated compressedsupercritical fluid, (d) expanding at least a portion of the first flowof heated compressed supercritical fluid in a first turbine so as toproduce a first flow of expanded supercritical fluid, with the expansionof the first flow of supercritical fluid generating shaft power, (e)returning the first flow of expanded supercritical fluid to the firstcompressor, (f) transferring heat from the first flow of expandedsupercritical fluid to the air prior to returning the first flow ofsupercritical fluid to the first compressor, (g) compressing a secondflow of supercritical fluid in a second compressor so as to produce asecond flow of compressed supercritical fluid, (h) transferring heatfrom the cooled combustion gas to the second flow of compressedsupercritical fluid so as to produce a second flow of heated compressedsupercritical fluid, (i) expanding the second flow of heated compressedsupercritical fluid in a second turbine so as to produce a second flowof expanded supercritical fluid and so as to generate additional shaftpower.

The invention also encompasses a system for generating shaft power usinga supercritical fluid cycle and an air cycle that comprises first andsecond flow paths. The first flow path directs the flow of a firstfluid, which comprises air, and comprises (i) a combustor connected tothe first flow path so as to receive at least a portion of the air, thecombustor supplied with a fossil fuel for combustion in the air, and inwhich the combustion of the fossil fuel in the air produces heatedcombustion gas, and (ii) a first turbine connected to the first flowpath. The second flow path directs the flow of a second fluid, whichcomprises a supercritical fluid, and that is separate from the firstflow path so as to prevent mixing of the air and the supercriticalfluids. The second flow path comprises (i) a first compressor connectedto the second flow path so as to receive the supercritical fluid forcompression therein and to discharge the compressed supercritical fluidinto the second flow path, and (ii) a second turbine for expansion ofthe supercritical fluid, with the second turbine connected to the secondflow path so as to discharge the expanded supercritical fluid into thesecond flow path. The system also comprises a first cross cycle heatexchanger connected to the first and second flow paths so as to (i)receive at least a portion of the air for transfer of heat thereto so asto heat the portion of the air prior to the portion of the air beingreceived by the combustor, and (ii) discharge the heated air into thefirst flow path, with the first cross cycle heat exchanger beingconnected to the second flow path so as to receive at least a portion ofthe expanded supercritical fluid discharged from the second turbine fortransfer of heat therefrom so as to cool at least the portion of theexpanded supercritical fluid, and to discharge the cooled expandedsupercritical fluid into the second flow path, with the expandedsupercritical fluid transferring heat to the air. The system alsoincludes a second cross cycle heat exchanger connected to the first andsecond flow paths so as to receive at least a portion of the combustiongas produced by the combustor for transfer of heat therefrom so as tocool the combustion gas, and discharge the cooled combustion gas intothe first flow path, and so as to receive at least a portion of thecompressed supercritical fluid from the first compressor for thetransfer of heat thereto so as to heat at least the portion of thecompressed supercritical fluid and discharge the heated supercriticalfluid into the second flow path, with the combustion gas transferringheat to the compressed supercritical fluid. In the system, the firstturbine is connected to the first flow path so as to receive at least aportion of the combustion gas produced by the combustor for expansiontherein, and to discharge the expanded combustion gas to the first flowpath, while the second turbine is connected to the second flow path soas to receive the heated supercritical fluid discharged from the secondcross cycle heat exchanger, the second turbine having a second shaft,with the expansion of the compressed supercritical fluid in the secondturbine driving rotation of the second shaft.

The invention also encompasses a system for generating shaft power usinga supercritical fluid cycle and an air cycle that comprises (a) acombustor for burning a fossil fuel in air so as to produce a combustiongas, (b) a first compressor for compressing a supercritical fluid so asto produce a compressed supercritical fluid, (c) a first cross cycleheat exchanger for transferring heat from the combustion gas to thecompressed supercritical fluid so as to produce a cooled combustion gasand a heated compressed supercritical fluid, (d) a first turbine forexpanding at least a portion of the heated compressed supercriticalfluid so as to produce an expanded supercritical fluid, with theexpansion of the supercritical fluid generating shaft power, (e) a flowpath for returning the expanded supercritical fluid to the firstcompressor, (f) a second cross cycle heat exchanger for transferringheat from the expanded supercritical fluid to the air so as to cool thesupercritical fluid to approximately its critical temperature prior toburning the fossil fuel in the air in the combustor and prior toreturning the supercritical fluid to the first compressor.

The invention also encompasses a system for generating shaft power usinga supercritical fluid cycle and an air cycle that comprises (a) acombustor for burning a fossil fuel in air so as to produce a combustiongas, (b) a first compressor for compressing a first flow ofsupercritical fluid so as to produce a first flow of compressedsupercritical fluid, (c) a first heat exchanger for transferring heatfrom the combustion gas to the first flow of the compressedsupercritical fluid so as to produce a cooled combustion gas and a firstflow of heated compressed supercritical fluid, (d) a first turbine forexpanding at least a portion of the first flow of heated compressedsupercritical fluid so as to produce a first flow of expandedsupercritical fluid, the expansion of the first flow of supercriticalfluid generating shaft power, (e) a flow path for returning the firstflow of expanded supercritical fluid to the first compressor, (f) asecond heat exchanger for transferring heat from the first flow ofexpanded supercritical fluid to the air prior to returning the firstflow of supercritical fluid to the first compressor, (g) a secondcompressor for compressing a second flow of supercritical fluid so as toproduce a second flow of compressed supercritical fluid, (h) a thirdheat exchanger for transferring heat from the cooled combustion gas tothe second flow of compressed supercritical fluid so as to produce asecond flow of heated compressed supercritical fluid, (i) a secondturbine for expanding the second flow of heated compressed supercriticalfluid so as to produce a second flow of expanded supercritical fluid andso as to generate additional shaft power.

The invention also encompasses a coupling from transmitting torque froma turbine shaft to a drive shaft in a system for generating shaft powerby expanding a supercritical fluid in the turbine. The couplingcomprising (a) an induction rotor adapted to be connected to the turbineshaft so as to rotate with the turbine shaft, (b) first and secondarmatures adapted to be connected to the drive shaft so as to rotatewith the drive shaft, (c) a magnet creating a magnetic flux within thecoupling, the magnet connected to the first and second armatures so asto rotate with the armatures, whereby rotation of the induction rotorimparts torque to the first and second armatures that causes rotation ofthe drive shaft, (d) a first flow path for directing a portion of thesupercritical fluid to the induction rotor for cooling the inductionrotor, whereby the portion of the supercritical fluid is heated, and (e)a second flow path for directing the heated supercritical fluid to theturbine for expansion therein. In one embodiment of the invention, thecoupling further comprises a pressure membrane disposed between theinduction rotor and the first and second armatures, with the pressuremembrane having an approximately spherically shaped surface.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic diagram of one embodiment of a power generationsystem according to the current invention in which the shaft powerdrives a turboprop.

FIG. 2 is a pressure-temperature phase diagram for supercritical carbondioxide in which the X-axis is temperature T and the y-axis is the logof pressure P.

FIG. 3 is a schematic diagram of an alternate embodiment of a powergeneration system according to the current invention.

FIG. 4 is a schematic diagram of another alternate embodiment of a powergeneration system according to the current invention.

FIG. 5 is a schematic diagram of portion of another alternate embodimentof a power generation system according to the current invention.

FIG. 6 is a schematic diagram of another alternate embodiment of a powergeneration system according to the current invention using reheat of thesupercritical fluid.

FIG. 7 is a schematic diagram of another alternate embodiment of a powergeneration system according to the current invention using reheat of thesupercritical fluid.

FIG. 8 is a schematic diagram of another alternate embodiment of a powergeneration system according to the current invention using reheat of thecombustion gas.

FIG. 9 is a schematic diagram of another alternate embodiment of a powergeneration system according to the current invention using reheat of thecombustion gas.

FIG. 10 is a schematic diagram of another alternate embodiment of apower generation system according to the current invention incorporatingsteam injection.

FIG. 11 is a schematic diagram of another alternate embodiment of apower generation system according to the current invention that alsogenerates hot water.

FIG. 12 is a schematic diagram of another alternate embodiment of apower generation system according to the current invention that alsogenerates hot water and uses a vacuum cycle.

FIG. 13 is a schematic diagram of another alternate embodiment of apower generation system according to the current invention thatincorporates a second supercritical fluid cycle.

FIG. 14 is a schematic diagram of another alternate embodiment of apower generation system according to the current invention thatincorporates a second supercritical fluid cycle.

FIG. 15 is a graph showing the change in specific heat c_(p) of SCO2 asa function of temperature in the vicinity of the critical temperatureT_(C).

FIG. 16 is a drawing, partially schematic, showing an apparatus formeasuring the temperature of SCO2 flowing to the inlet of the SCO2compressor.

FIG. 17 is a graph of the speed of sound, in m/s, in SCO2 as a functionof temperature, in ° K, at 7.4 MPa.

FIG. 18 is a drawing, partially schematic, showing another apparatus formeasuring the temperature of SCO2 flowing to the inlet of the SCO2compressor.

FIG. 19 is a longitudinal cross section through the turbine couplingportion of a power generation system according to the current invention.

FIG. 20 is a transverse cross section through the turbine coupling shownin FIG. 19, taken along line XX-XX.

FIGS. 21 and 22 are isometric view of cross sections through the turbinecoupling shown in FIG. 19.

FIG. 23 is a cross section through a portion of the turbine couplingshown in FIG. 19.

DETAILED DESCRIPTION OF ILLUSTRATIVE EMBODIMENTS

One embodiment of a fossil fuel fired, dual cycle, supercriticalfluid-air system for generating shaft power according to the currentinvention is shown in FIG. 1. The system comprises a first Brayton cyclesystem 2, in which the working fluid is a supercritical fluid, such assupercritical carbon dioxide (SCO2), and a second Brayton cycle system 4in which the working fluid is ambient air. Each of these cyclescomprises a flow path 6 and 23, which may be formed by piping, ductworkor other conduits as appropriate, to which various components, such ascompressors, turbines, combustors and heat exchangers, are connected.The SCO2 cycle flow path 6 and air breathing cycle flow path 23 arepreferably separate so that little or no mixing occurs between thefluids in the two flow paths.

The supercritical Brayton cycle system 2 forms a closed cycle flow path6 through which the supercritical fluid flows. Initially, a stream 3 ofsupercritical fluid is supplied to the inlet of a compressor 8, whichmay be an axial, radial or even reciprocating type. A flow meter 32measures the flow rate of the fluid supplied to the compressor inlet.This provides a means for inventory control of total SCO2 mass in theclosed system as well as for control of transient flow behavior.Preferably, the supercritical fluid enters the inlet of the compressor 8after it has been cooled and expanded, as discussed below, to atemperature and pressure that is close to its critical point. Thiscritical point is illustrated in FIG. 2, which is a pressure-temperaturephase diagram for a supercritical fluid, in this case, carbon dioxide.The carbon dioxide is a solid in region A, a liquid in region B and agas in region C. In the region D above the temperature and pressure atthe critical point E, the carbon dioxide exists as a supercriticalfluid. Thus, as used herein, the term “supercritical fluid” refers to afluid in which distinct liquid and gaseous phases do not exist, and term“critical point” of a supercritical fluid refers to the lowesttemperature and pressure at which the substance can be said to be in asupercritical state. The terms “critical temperature” and “criticalpressure” refer to the temperature and pressure at the critical point.For carbon dioxide, the critical point is approximately 304.2° K and7.35 MPa. Preferably, the supercritical fluid entering the compressor 8is cooled to within at least ±2° K of its critical point, morepreferably to within ±1° K of its critical point, and most preferably towithin ±0.2° K of its critical point.

After compression in the compressor 8, the stream 5 of SCO2 is heated ina cross cycle heat exchanger 10, which may be a Printed Circuit HeatExchanger (PCHE) or other type as appropriate and which is connected tothe flow paths 6 and 23 of both the SCO2 and air breathing cycles. Asused herein, the term “cross cycle heat exchanger” refers to a heatexchanger that receives both air or combustion gas from the airbreathing cycle as well as a supercritical fluid from the supercriticalfluid cycle and transfers heat between the fluids in the two cycles. Thestream 7 of heated SCO2 from the heat exchanger 10 is then directed tothe inlet of a turbine 12, which may be an axial, radial or mixed flowtype, in which the SCO2 is expanded and produces shaft power that drivesboth the SCO2 compressor 8, via shaft 9, and a turboprop 14, via a shaft17 and a reduction gear 16. After expansion in the turbine 12, thestream 9 of SCO2 is cooled in a second cross cycle heat exchanger 18,which may be a PCHE type and which is connected to the flow paths 6 and23 of both the SCO2 and air breathing cycles. The stream 3 of cooledSCO2 is returned to the inlet of the compressor 8 via the flow path 6.Preferably the cross cycle heat exchanger 18 has sufficient surface areato cool the SCO2 returned to the compressor 8 to a temperature close toits critical temperature as discussed above. Additional SCO2 from asupply 31 can be introduced into the stream 3 of SCO2 directed to thecompressor 8 to make up for any leakage of SCO2 from the system. Inaddition, the introduction of additional SCO2 into the system can bemodulated to attenuate system dynamics during transients. In any event,the SCO2 3 is returned to the inlet of the compressor 8 and the steps ofcompressing-heating-expanding-cooling are repeated.

A shown in FIG. 1, the air breathing Brayton system 4 portion of theoverall system forms an open flow path 23. Initially, ambient air 11 issupplied to a compressor 20 which may be axial, radial or reciprocatingtype. The stream 13 of compressed air from the compressor 20 is thenheated in the heat exchanger 18 by the transfer of heat from the SCO2after the SCO2 has been expanded in the turbine 12. The stream 15 ofheated compressed air is then directed to a combustor 24 into which afossil fuel 27, such as jet fuel, diesel fuel, natural gas or bio-fuel,is introduced by a fuel controller 28 and combusted in the air so as toproduce hot combustion gas. The stream 37 of the combustion gas from thecombustor 24 is directed to the heat exchanger 10 where heat istransferred to the SCO2, as discussed above. After exiting the heatexchanger 10, the stream 19 of combustion gas is expanded in a turbine26 which may be an axial, radial or mixed flow type, and which producespower to drive the air compressor 20, via shaft 21. After expansion inthe turbine 26, the combustion gas 47 is exhausted to atmosphere.

The operation of the system shown in FIG. 1 will now be illustrated byway of an example of predicted results. In this example which is for aturboprop/turbofan application, the ambient air, which is at standardday conditions at 9000 m, is supplied to the inlet of the compressor 20at 229.7° K and 32 KPa. The air compressor 20 is operated at acompression ratio of only about 2.0 so that the compressed airdischarged by the compressor and directed to the heat exchanger 18 is ata temperature and pressure of only about 295° K and 65 KPa. The SCO2exhausted from the turbine 12 and directed to the heat exchanger 18 isat a temperature and pressure of about 935° K and 7.5 Mpa. The heatexchanger 18 has sufficient heat transfer surface area so that thecompressed air is heated from about 295° K to about 923° K and the SCO2is cooled from about 935° K to about 305° K, close to its criticaltemperature. In order to control the temperature of the SCO2 enteringthe compressor 8, so as to maintain it close to its criticaltemperature, the compressed air discharged by the compressor 20 can bedirected to a heat exchanger (not shown) supplied with a cooling fluid.The flow rate and/or temperature of the cooling fluid can be varied soas to adjust the temperature of the compressed air entering the heatexchanger 18 so that the heat transfer in heat exchanger 18 cools theSCO2 to a temperature close to its critical temperature.

In this example, sufficient fuel is burned in the combustor 24 to heatthe compressed air discharged from the heat exchanger 18 from about 886°K to about 1121° K, the temperature at which it enters the heatexchanger 10. The SCO2 compressor 8 operates at a much highercompression ratio than the air breathing compressor 20, and compressesthe SCO2 from its incoming pressure of 7.4 Mpa, close to the criticalpressure, to a pressure of approximately 25.9 MPa and a temperature ofapproximately 383° K, the temperature and pressure at which the SCO2 issupplied to the heat exchanger 10. Heat exchanger 10 contains sufficientheat transfer surface area so that the SCO2 is heated from about 383° Kto about 1103° K and the combustion gases are cooled from about 1121° Kto about 399° K. After the combustion gas is expanded in the turbine 26,it is exhausted to atmosphere at about 341° K. After the heated SCO2 isexpanded in the turbine 12, it is exhausted at about 935° K to heatexchanger 18, where it is cooled to about 305° K prior to return to theinlet of the SCO2 compressor 8, as discussed above.

The system illustrated in FIG. 1 has several important advantages.Supercritical fluids such as SCO2 have a very low dynamic viscosity andhigh specific heat, which facilitates the use of heat exchangers thathave a low pressure drop and high effectiveness for a given size andweight. Further, since a supercritical fluid, such as SCO2, has adensity and compressibility approaching that of a liquid, the workrequired to compress the fluid to the desired pressure ratio is muchlower than it would be for an ideal gas such as air. This not onlyincreases the net work available from the SCO2 turbine 12, it results inlower compressor discharge temperature from the SCO2 compressor which,in turn, increases the heat transfer from the combustion gas to the SCO2discharged by the compressor 8 that is achieved in the heat exchanger10.

Moreover, the air compressor 20 is operated at a relatively low pressureratio so that the air discharging from the air compressor is at arelatively low temperature (295° K in the example above), therebyincreasing the heat that can be recovered from the SCO2 in the heatexchanger 18. As a result of the high heat transfer in heat exchanger18, it may be unnecessary to employ any “external” cooler to cool theSCO2 exhausted from the turbine 12 to the appropriatetemperature—preferably close to its critical temperature—for return tothe inlet of the compressor 8. Thus, the cycle rejection heat that wouldotherwise be lost from the cycle to an external heat sink, such ascooling water from a cooling tower, in order to cool the SCO2 followingexpansion in the turbine 12 is retained within the system.

Assuming an efficiency of 87% for the turbines 12 and 26 andefficiencies of 85% and 87%, respectively, for the SCO2 compressor 8 andthe air compressor 20, the overall cycle efficiency of the system shownin FIG. 1, and operated as discussed above, is calculated to be about54%.

Although it is not necessary to incorporate all of the followingcharacteristics in order to achieve the benefits of the currentinvention, the characteristics of the system shown in FIG. 1 thatcontribute to high thermal efficiency are the use of (i) two Braytoncycles operating in parallel, one of which (system 2) operates on asupercritical fluid, preferably SCO2, and the other of which (system 4)operates on ambient air, which acts similarly to an ideal gas, (ii) afirst cross cycle heat exchanger (heat exchanger 18) in which cyclerejection heat which is normally lost in cooling the SCO2 to the desiredcompressor inlet temperature (preferably close to its criticaltemperature) is instead transferred from the SCO2 exhausted from theturbine 12 in the SCO2 cycle to the air discharging from the compressor20 in the air breathing cycle, (iii) a second cross cycle heat exchange(heat exchanger 10) in which heat is transferred from the combustion gasof the air breathing cycle to the compressor discharge in the SCO2cycle, and (iv) a relatively low pressure ratio in the air breathingcompressor 20, which would be considered suboptimal according toconvention thinking but which, in the current invention, allows thereturn of a large amount of heat to the system in cooling the SCO2exhausted from the turbine to the desired compressor inlet temperature.

Although the system discussed above has been illustrated as supplyingshaft power for a turboprop, it should be understood that the inventionis also applicable to any other application utilizing shaft power,including but not limited to electrical power generation, navalpropulsion systems, rail engine drives, hybrid drives for automobilesand trucks, gas booster pumps for the oil and gas industry, agriculturalpumping applications, and construction equipment drives.

Although optimal benefits are obtained according to the currentinvention by using both a supercritical Brayton cycle system and an airbreathing Brayton cycle system in tandem, certain benefits cannevertheless be achieved by operating both systems illustrated in FIG. 1on ambient air. Such an embodiment provides the ability to use a varietyof fuels, including low quality fossil fuels, biomass and even solarpower. In such a system, the products of combustion, including ash,would not pass through the higher temperature turbine 12 and thereforecause no fouling of turbine air cooled components. Ash and combustionbyproducts would pass through the low expansion turbine 26 at a muchlower temperature such that cooling passages are unnecessary and makinga particulate resistant turbine fairly straightforward. Particulatebuild-up in the downstream heat exchanger 10 could be dealt with throughperiodic wash or cleaning cycles, which could be done during routineshutdowns.

Although the embodiment in FIG. 1 has been depicted as directing thecombustion gas 37 from the combustor 24 to the heat exchanger 10 andthen to the turbine 26, the invention could also be practiced bydirecting the combustion gas from the combustor to the turbine 26 first,for expansion therein, and then directing the expanded gas to the heatexchanger 10, as discussed below in connection with the embodiment shownin FIG. 4, for example.

Another embodiment of the current invention is shown in FIG. 3, withsimilar components identified by similar reference numerals. In thisembodiment, a cooler 22, supplied with a cooling fluid 30, such as waterfrom a cooling tower in a land based application or air cooler in anaviation application, is used to control the temperature of the air 13′discharging from the air compressor 20 prior to its introduction intoheat exchanger 18 so as to control the temperature of the SCO2 enteringthe SCO2 compressor 8. As previously discussed, preferably, thetemperature of the SCO2 entering the compressor 8 is controlled to closeto its critical temperature. This temperature control may be effected bycontrolling the flow rate and/or temperature of cooling fluid suppliedto the cooler 22.

In the FIG. 3 embodiment, two SCO2 turbines are utilized. The firstturbine 12′ is a SCO2 compressor turbine that drives the SCO2 compressor8, while the second SCO2 turbine 12″ is a power turbine that providesthe power output of the system. Further, in this embodiment, the flow 41of SCO2 exhausting from the SCO2 compressor turbine 12′ can be dividedby a valve 38 into two streams. A first stream 42 can be directed forexpansion in the power turbine 12″, while a second stream 44 can bedirected to an isenthalpic expansion nozzle 34 that reduces the pressureof the SCO2 to close to that of the inlet pressure of the SCO2compressor 8. These two streams are joined at junction 40 and thecombined stream directed to heat exchanger 18 as before. Note thatalthough the nozzle 34 is shown in FIG. 3 as being utilized so that aportion of the SCO2 bypasses the second SCO2 turbine 12″, the nozzle 34could also be incorporated into an embodiment like that shown in FIG. 1in which only one SCO2 turbine is utilized so that a portion of the SCO2directed to the nozzle bypassed that single SCO2 turbine.

Although it may decrease efficiency, diverting a portion 44 of the SCO2discharged from the SCO2 compressor turbine 12′ so that it bypasses thepower turbine 12″ allows the system to maintain optimum efficiency whenthere is little or no load on the power turbine 12″ by maintaining theturbine 12′ at its design point inlet temperature and pressure ratio.This not only increases the efficiency of the power turbine 12″ butreduces the deleterious effects of thermal cycling on the useful life ofthe “hot” turbine components. In addition, operation of the valve 38enables the power turbine 12″ to quickly respond to an increase in powerdemand, and increases the stability of the system in the face of powerdemand transients. Advantageously, although the pressure of the SCO2 isreduced in the isenthalpic expansion nozzle 34, its temperature remainshigh so that the unused heat is returned to the system in heat exchanger18, increasing the temperature of the air directed to the combustor 24and, therefore, reducing the fuel that must be burned to achieve thedesired combustor outlet temperature. Note that although the expansionnozzle 34 is depicted in FIG. 3 as receiving a stream 42 of partiallyexpanded SCO2 from the first turbine 12′ of a system employing two SCO2turbines 12′ and 12″, the expansion nozzle 34 could also beincorporation into a system employing only one SCO2 turbine 12, such asthat shown in FIG. 1, so that a portion of the SCO2 from the heatexchanger 10 bypassed the turbine 12 and was directed, after expansion,to the heat exchanger 18.

Another refinement in the system shown in FIG. 3 is the use of two heatexchangers 10′ and 10″ to transfer heat from the combustion gas to theSCO2 discharging from the SCO2 compressor 8. Since the temperatures ofthe SCO2 and combustion gas in the heat exchanger 10″ are lower than inheat exchanger 10′, this arrangement allows less expensive materials tobe used in the heat exchanger 10″.

As discussed further below, according to one embodiment of theinvention, an eddy current coupling 36 is used to transmit power fromthe power turbine shaft 17 to the driven shaft 58, which may be theshaft of a turboprop or an electrical generator, for example.Consequently, a portion 52 of the SCO2 stream 48 discharging from theSCO2 compressor 8 is directed by a valve 51 to the eddy current couplingfor cooling purposes, while the remaining portion 50 of SCO2 stream 48is directed to heat exchanger 10″. Preferably, after absorbing heat inthe eddy current coupling 36, the stream 54 of heated SCO2 is directedto heat exchanger 10′, where it mixes with stream 53, for furtherheating and then expansion in the SCO2 turbines so that the heatabsorbed from the coupling is not lost from the system.

FIG. 4 illustrates a variation on the FIG. 3 embodiment that may beparticularly useful in land based applications, in which the compressionratio in the air compressor 20 must be kept low in order to maintain alower compressor discharger temperature so as to achieve maximum heattransfer in heat exchanger 18. In this embodiment, the combustion gasdischarged from the combustor 24 is first expanded in the air turbine 26and then subsequently directed to the heat exchangers 10′ and 10″ fortransfer of heat to the SCO2 discharged by the SCO2 compressor 8. Afterflowing through the heat exchangers 10′ and 10″, the cooled combustiongas is exhausted to atmosphere.

FIG. 5 shows a portion of a system similar to the FIGS. 3 and 4embodiments—the portions of the system not shown in FIG. 5 are the sameas those in FIGS. 3 and 4. The embodiment shown in FIG. 5 allows theflexibility of operation according to either the FIG. 3 or FIG. 4embodiments, or a combination of the two embodiments, so as tofacilitate maximum performance either at high altitudes or sea level. Inthis embodiment, a first valve 62 is incorporated into the flow pathdownstream of the combustor 24, a second valve 68 is incorporateddownstream of the air turbine 26, and a third valve 77 is incorporateddownstream of the heat exchanger 10″. Operation of valve 62 allows allor a portion of the combustion gases 19 discharging from the combustor24 to be initially directed either to the air turbine 26 (stream 64) orto the heat exchanger 10′ (stream 74). Operation of valve 68 allows allor a portion of the gas 66 exhausting from the air turbine 26 to bedirected either to the heat exchanger 10′ (stream 72) or exhausted toatmosphere (stream 70). Operation of valve 77 allows all or a portion ofthe gas discharged from heat exchanger 10″ to be directed either to theair turbine 26 (stream 80) or exhausted to atmosphere (stream 78). Theflow splits provided by the valves 62, 68 and 74 can be adjustedcontinuously in order to achieve peak efficiency depending onatmospheric conditions.

In general, expanding the combustion gas in the turbine 26 beforecooling it in the cross cycle heat exchangers 10′ and 10″ providesimproved performance in land based applications, and at sea level or lowaltitude in aviation applications, whereas cooling the combustion gas inthe cross cycle heat exchangers before expanding it in the turbineyields better results at altitude.

FIG. 6 illustrates an embodiment of the invention that is similar toFIG. 4 but which incorporates reheating of the SCO2. In this embodiment,the stream 5 of compressed SCO2 discharged from the SCO2 compressor 8 isdirected to heat exchangers 100 and 101, where it is heated by thetransfer of heat from the combustion gas 104 exhausted from the airturbine 26 and then expanded in the SCO2 compressor turbine 12′, asbefore. However, after being expanded in the SCO2 compressor turbine12′, the stream 102 of partially expanded SCO2 is again directed to heatexchanger 100, where it is reheated by the transfer of heat from thecombustion gas 104 exhausted from the air turbine 26, thereby reheatingthe SCO2. From the heat exchanger 100, the stream 106 of reheated SCO2is directed to the splitter valve 38, as before, so that, if desired,the flow can be divided between a first stream 108 that is expanded inthe power turbine 12″ and a second stream 110 that is expanded in theisenthalpic expansion nozzle 34. The streams 112 and 114 of expandedSCO2 are then combined as stream 116 and directed to the heat exchanger18 where the SCO2 is cooled by the transfer of heat to the airdischarging from the air compressor 20. Reheating the SCO2 afterexpansion in the SCO2 compressor turbine 12′ but prior to expansion inthe power turbine 12″ in this embodiment has the advantage of increasingthe overall thermal efficiency of the system by an expected amount ofabout 2%. In this embodiment, power from the power turbine shaft 17drives an electric generator 90. However, it should be understood thatthis system, as well as the systems described below, can be used in anyapplication requiring the use of shaft power.

The embodiment illustrated in FIG. 7 is similar to that of FIG. 3 exceptthat the combustion gas 120 from the combustor 24 heats both thecompressed SCO2 from the SCO2 compressor 8 and reheats the SCO2 124exhausted from the SCO2 compressor turbine 12′ in heat exchanger 100 asin the FIG. 6 embodiment.

As discussed in connection with the embodiment shown in FIG. 5, valvescould be incorporated into the flow path of the embodiments shown inFIG. 6 or 7 so that operation could be shifted from that of FIG. 6, inwhich combustion gases are expanded in the air turbine 26 before beingdirected to the heat exchangers 100 and 101, to that of FIG. 7, in whichcombustion gases are directed to heat exchangers 100 and 101 beforebeing expanded in the air turbine 26. Alternatively, the valves could beoperated so that the system operated in both modes simultaneously, withthe split between the two modes being varied as necessary to achieveoptimum performance.

In the embodiment illustrated in FIG. 8, which bears similarities tothat of the FIG. 6 embodiment, the stream 5 of compressed SCO2discharged by the SCO2 compressor 8 is sequentially heated in heatexchangers 150 and 130, and the heated SCO2 is then expanded in the SCO2compressor turbine 12′, as before. In heat exchanger 130, heat istransferred to the SCO2 from the combustion gas 148 exhausted from theair turbine 26, as before. However, in this embodiment, the partiallycooled combustion gas 152 exiting the heat exchanger 130 is reheated inreheat combustor 140 by burning additional fuel in the combustion gasunder the operation of a fuel control 142. From the reheat combustor 140the reheated combustion gas 146 is then directed to heat exchanger 144,which heats the SCO2 discharged from the SCO2 compressor turbine 12′prior to its expansion in the power turbine 12″ (or its expansion in thenozzle 34). The combustion gas 154 discharged from heat exchanger 144 isthen directed to heat exchanger 150, in which it transfers heat to theSCO2 compressor discharge 5, and the combustion gas 156 is thenexhausted to atmosphere. This arrangement has the benefit again ofincreasing the overall thermal efficiency of a system without reheat byan expected amount of about 2%.

In the embodiment illustrated in FIG. 9, which bears similarities tothat of the FIG. 7 embodiment, the stream 5 of compressed SCO2discharged by the SCO2 compressor 8 is sequentially heated in heatexchangers 150 and 130, and the heated SCO2 is then expanded in the SCO2compressor turbine 12′, as before. In heat exchanger 130, heat istransferred to the SCO2 from the combustion gas 176 from the combustor24, as before. However, the partially cooled combustion gas 178 exitingthe heat exchanger 130 is then reheated in reheat combustor 140 byburning additional fuel in the combustion gas under the operation of afuel control 142. From the reheat combustor 144, the reheated combustiongas 180 is directed to heat exchanger 144, which heats the SCO2discharged from the SCO2 compressor turbine 12′ prior to its expansionin the power turbine 12″ (or its expansion in the nozzle 34). Thecombustion gas 182 is then directed to heat exchanger 150, in which ittransfers heat to the SCO2 compressor discharge 5, and the combustiongas 184 is then expanded in the air turbine 26, after which thecombustion gas 186 is exhausted to atmosphere. This arrangement has thebenefit of raising the overall thermal efficiency of a system withoutreheat by an expected amount of about 2%.

As discussed in connection with the embodiment shown in FIG. 5, valvescould be incorporated into the flow path of the embodiments shown inFIG. 8 or 9 so that operation could be shifted from that of FIG. 8 tothat of FIG. 9. Alternatively, the valves could be operated so that thesystem operated in both modes simultaneously, with the split between thetwo modes being varied as necessary to achieve optimum performance.

The embodiment shown in FIG. 10 is similar to that shown in FIG. 9except that the combustion gas 202 expanded in the air turbine 26 isdirected to a water boiler 200 before being exhausted to atmosphere. Thewater boiler 200 transfers heat from the combustion gas to water 206,thereby generating steam 208. The steam 208 is directed to a heatexchanger 210 where it is superheated by the transfer of heat from thestream 204 of SCO2 expanded by the power turbine 12″ (or expansionnozzle 34) prior to introduction into the heat exchanger 18. Thesuperheated steam 212 is then injected into the combustor 24 along withthe fuel, thereby increasing the mass flow of the combustion gas 216directed through the heat exchangers 130, 144 and 150 and the airturbine 26. Although the injection of the steam 212 into the combustor24 increases the fuel required to achieve a given combustor outlettemperature, since the additional heat recovered from the air turbineexhaust gas 202 by the water boiler 200 is returned to the cycle, theefficiency is increased. The injection of steam in the combustor canalso reduce the generation of NOx, a pollutant.

FIG. 11 illustrates an embodiment of the invention in which bothelectrical power and hot water, for example for district heating, aregenerated. In this embodiment, heat is transferred to ambient water 228supplied to the airside cooler 22 to lower the temperature of thecompressed air 13′ discharging from the air compressor 20 prior to itsintroduction into the heat exchanger 18, as previously discussed. Theslightly heated water 230 discharged from the airside cooler 22 is thendirected to an SCO2 intercooler 220 positioned between first and secondSCO2 compressors 8′ and 8″ connected in series. Partially compressedSCO2 222 is directed from the first compressor 8′ to the SCO2intercooler 220 in which heat is transferred from the partiallycompressed SCO2 to the incoming water 230. After cooling, the cooledSCO2 224 discharged from the first compressor 8′ is then directed to thesecond compressor 8″, where it undergoes compression to its desiredvalue. In this embodiment, the heated water 232 discharged from the SCO2intercooler 220 can advantageously be used for heating purposes.

Note that the benefit of compressor interstage cooling is well known asit reduces the amount of work required to achieve a desired pressureratio at the compressor discharge. In the embodiment shown in FIG. 11,the SCO2 intercooler 220 is designed to reduce the temperature of theinterstage SCO2 222 to just above its supercritical temperature, whichresults in an expected reduction of nearly 25% in the power required toachieve the desired pressure. Further, the characteristic intercoolerinlet and outlet temperatures of the SCO2 lend themselves to heatingwater as part of a combined heat and power implementation. Heating ofthe water 228 in the airside cooler 22 raises the temperature of thewater only slightly since, in one embodiment for example, the heattransferred in the airside cooler 22 is only about 13 KW per Kg/s ofSCO2 mass flow. By varying the percentage of work done by the SCO2compressors 8′ and 8″ (in other words varying the ratio of pressureratios) the amount of heat extracted in the SCO2 intercooler 220 can bevaried from, for example, 20 KW per Kg/s to about 150 KW per Kg/s. Thus,the magnitude of the heat transferred to water can be about the same asthe heat transfer generating power. This differs substantially fromconventional combined heat and power systems since they generallygenerate about two times as much heat as electricity. In one embodimentof the system shown in FIG. 11, the water 230 discharged from theairside cooler 22 flows to the SCO2 intercooler 220, where up to about150 KW is extracted in order to lower the temperature of the SCO2 toclose to its critical temperature of about 305° K. The heat transfer inthe SCO2 intercooler 220 raises the temperature of the water 232discharged from the SCO2 in cooler 220 to about 160° F., which is quitesuitable for heating and cooling (trigeneration) applications.

Note too that the SCO2 intercooler 220 results in lowering thetemperature of the stream 226 of SCO2 discharged by the compressor 8″,which would require an increase in the amount of heat input to thecompressor discharge stream, and by implication, the amount of fuelburned in the combustor 24 to achieve the desired inlet temperature inthe SCO2 turbine 12′. However, in this case, the heat source for theSCO2 is the flow of combustion gases from the combustor 24 so thereduced SCO2 compressor discharge temperature merely results in areduction in the temperature of the combustion gas exhausted toatmosphere from the air turbine 26, requiring little to no increase infuel flow to the combustor 24. The embodiment illustrated in FIG. 11 cansimultaneously provide hot water for heating applications pluselectricity at high efficiency so that an overall thermal efficiency onthe order of 90% is expected.

Although the embodiment in FIG. 11 has been depicted in a system inwhich the combustion gas from the combustor 24 is first directed to theturbine 26 for expansion therein and then directed to the heat exchanger130, the invention could also be practiced by directing the combustiongas from the combustor to the heat exchanger 130 first and thendirecting the combustion gas to the turbine 26 as shown, for example, inthe embodiment shown in FIG. 3.

FIG. 12 illustrates another embodiment of the invention applied to acombined heat and power system making use of a vacuum cycle. Ambientair—that is, air at ambient temperature and pressure—300 is drawn into across cycle heat exchanger 316 in which it absorbs heat from theexpanded SCO2 318 discharged from the SCO2 power turbine 12″, therebycooling the SCO2 317 directed to the SCO2 compressor 8 to close to itcritical temperature. The heated air 301 from the heat exchanger 316 isthen further heated in the combustor 302 by burning a fossil fuel (notshown). The resulting combustion gas 303 is then expanded in a turbine304 to below atmospheric pressure and the expanded gas 305 is directedto cross cycle heat exchangers 306 and 308 where it transfers heat tothe compressed SCO2 322 discharged from the SCO2 compressor 8. Althoughtwo cross cycle heat exchangers in series are shown in FIG. 12, theinvention could also be practiced using a single heat exchanger or morethan two heat exchangers in series. The heated SCO2 320 is then expandedin the turbines 12′ and 12″ so as to generate shaft power to drive theSCO2 compressor 8 and, for example, an electric generator 90, aspreviously discussed. Since this embodiment, as well as the embodimentshown in FIG. 13 discussed below, uses ambient air 300 as the coolingfluid in the cross cycle heat exchanger 316, rather than compressordischarge air, to cool the SCO2 317 returned to the compressor inlet, itis not necessary to use a cooler, such as cooler 22 shown in FIG. 11, tocool the air directed to the cross cycle heat exchanger, therebyavoiding the loss of heat from the cycle.

From the heat exchangers 306 and 308 the cooled combustion gas 309 isdirected to a water heater 310 supplied with water 311, which may be atambient temperature. In the water heater 310, heat is transferred fromthe combustion gas 309 to the water 311 so as to discharge heated water315. The heated water may be advantageously used for district heating,for example, or for any application making use of heated water. Thecooled combustion gas 312 discharged from the water heater 310 isdirected to a compressor 313 that increases the pressure of thecombustion gas above that of atmospheric pressure so that the combustiongas 314 can be exhausted to atmosphere.

FIG. 13 illustrates another embodiment of the invention applied to acombined heat and power system making use of a vacuum cycle along withtwo SCO2 cycles. Ambient air 300 is drawn into a cross cycle heatexchanger 316 in which it absorbs heat from the expanded SCO2 318discharged from the SCO2 power turbine 12″, thereby cooling the SCO2 317directed to the SCO2 compressor 8 to close to it critical temperature,as before. The heated air 301 from the heat exchanger 316 is thenfurther heated in the combustor 302 by burning a fossil fuel (notshown), as before. The resulting combustion gas 303 is then directed tothe cross cycle heat exchangers 306 and 308, rather than to a turbine asin the FIG. 12 embodiment. In the cross cycle heat exchangers 306 and308, heat is transferred from the combustion gas 303 to the compressedSCO2 322 discharged from the SCO2 compressor 8. The heated SCO2 320 isthen expanded in the turbines 12′ and 12″ so as to generate shaft powerto drive the SCO2 compressor 8 and, for example, an electric generator90, as previously discussed.

From the cross cycle heat exchangers 306 and 308, the partially cooledcombustion gas 341 transfers heat to a second SCO2 cycle through which asecond stream of SCO2 flows. In particular, the combustion gas 341 isdirected to a secondary cross cycle heat exchanger 336 where it isfurther cooled by transferring heat to SCO2 335 discharged from asecondary SCO2 compressor 334. The further cooled combustion gas 342 isthen directed to a compressor 313. As a result of the pressure dropthrough the heat exchangers, the combustion gas at the compressor inletwill be sub atmospheric. The compressor 313 increases the pressure ofthe combustion gas above that of atmospheric pressure so that thecombustion gas 314 can be exhausted to atmosphere.

The heated SCO2 337 discharged from the secondary cross cycle heatexchanger 336 is expanded in a secondary SCO2 turbine 330, whichgenerates shaft power to drive the secondary SCO2 compressor 334. Theexpanded SCO2 331 discharged from the turbine 330 is then directed to awater heater 395, where it transfers heat to water 311, thereby coolingthe SCO2 333 to close to its critical temperature before it is returnedto the secondary SCO2 compressor 334. The heated water 315 mayadvantageously be used for district heating, for example, as previouslydiscussed.

FIG. 14 illustrates yet another embodiment of the invention applied to acombined heat and power system making use of a vacuum cycle along withtwo SCO2 cycles. Ambient air 300 is drawn into a compressor 370 and thecompressed air is then cooled in a cooler 22 by transferring heat towater (not shown) as in the FIG. 11 embodiment. The cooled compressedair 373 is directed to a cross cycle heat exchanger 316, where heat istransferred to it from the SCO2 so as to cool the SCO2 317 directed tothe SCO2 compressor 8 to close to its critical temperature.

The heated air 301 from the heat exchanger 316 is then further heated inthe combustor 302 by burning a fossil fuel (not shown), as before. Theresulting combustion gas 303 is then directed to the cross cycle heatexchangers 306 and 308 in which heat is transferred from the combustiongas 303 to the compressed SCO2 322 discharged from the SCO2 compressor8, as in the FIG. 13 embodiment. The heated SCO2 320 is then expanded inthe turbines 12′ and 12″ so as to generate shaft power to drive the SCO2compressor 8 and, for example, an electric generator 90, as previouslydiscussed.

From the cross cycle heat exchangers 306 and 308, the partially cooledcombustion gas 341 transfers heat to a second SCO2 cycle, as in the FIG.13 embodiment. In particular, the combustion gas 341 is directed to asecondary cross cycle heat exchanger 336 where it is further cooled bytransferring heat to SCO2 335 discharged from a secondary SCO2compressor 334. The further cooled combustion gas 342 is then exhaustedto atmosphere.

The heated SCO2 337 discharged from the secondary cross cycle heatexchanger 336 is expanded in a secondary SCO2 turbine 330, whichgenerates shaft power to drive the secondary SCO2 compressor 334 as wellas the air compressor 370. The expanded SCO2 331 discharged from theturbine 330 is then directed to a water heater 395, where it transfersheat to water 311, thereby cooling the SCO2 333 to close to its criticaltemperature before it is returned to the secondary SCO2 compressor 334.The heated water 315 may advantageously be used for district heating,for example, as previously discussed.

It can be noted that whereas in the FIG. 13 embodiment, the aircompressor 313 “pulls” the air through the system, in the FIG. 14embodiment, the air compressor 370 “pushes” the air through the system.Also note that in both the FIGS. 13 and 14 embodiments, the air cycleportion of the system includes a compressor and combustor but noturbine.

As shown in FIG. 15, as is typical for supercritical fluids, thespecific heat of SCO2 changes dramatically around its criticaltemperature. Therefore, as previously discussed, it is important tomaintain the temperature of the SCO2 at the SCO2 compressor inlet asclose as possible to the critical temperature. In fact, it has beenfound that the thermal efficiency of the fossil fuel fired, dual cycle,supercritical fluid-air systems described herein can change in the orderof a few percent as a result of a change in the temperature of the SCO2at the inlet to the SCO2 compressor of only a few degrees Kelvin.Unfortunately, thermocouples typically used to measure temperature ingas turbine system are typically accurate to only a few degrees Kelvin.Consequently, according to one aspect of the current invention, methodsare provided for more accurately measuring the temperature of the SCO2at the SCO2 compressor inlet.

FIG. 16 shows one embodiment of an apparatus for measuring thetemperature of the SCO2 entering the SCO2 compressor inlet according tothe current invention. A bypass conduit 402 is connected to the mainconduit 400 that directs the stream 412 of SCO2 to the inlet of the SCO2compressor. A bypass stream 414 of SCO2 flows through the bypass conduit402. A pressure sensor 404, such as a piezo-electric type or other asappropriate, is incorporated into the bypass conduit 402 to measure thestatic pressure of the SCO2 in the bypass conduit. Upstream anddownstream temperature sensors 406 and 408, respectively, which may bethermocouples or other types of temperature sensors, are installed oneither side of a heat source 410. The heat source 410, such as anelectric coil or ceramic heater, introduces a known amount of heat intothe bypass stream 414 of SCO2. Preferably, the temperature sensors 406and 408 are spaced approximately ½ m apart.

By measuring the temperature of the SCO2 at both temperature sensors 406and 408 simultaneously when no heat is generated by the heat source 410,so that both sensors are measuring the same total temperature, thetemperature sensors can be corrected to account for deviations betweenthe two. The temperature measurements are then repeated while a knownamount of heat is being introduced into the SCO2 stream by the heatsource 410. The specific heat of the SCO2 can be determined by comparingthe increase in temperature between temperature sensors 406 and 408,taking into account the mass flow rate of the SCO2 through the conduit402, which can be inferred by analysis. This specific heat can then becompared to data for specific heat versus temperature at the staticpressure measured by the sensor 404 to accurately determine thetemperature of the SCO2 flowing in the main conduit 400.

As shown in FIG. 17, the speed of sound in SCO2 varies dramaticallyaround the critical point. As a result, the temperature of SCO2 can alsobe determined by calculating the speed of sound in the fluid andcross-referencing the measured speed of sound to the temperature of SCO2as a function of pressure and the speed of sound. Accordingly, anotherapparatus for measuring the temperature of the SCO2 directed to theinlet of the SCO2 compressor is shown in FIG. 18. A conduit 500 carriesthe stream 502 of SCO2 directed to the inlet of the SCO2 compressor. Apressure sensor, such as pressure sensor 404 shown in FIG. 16, is usedto measure the pressure of the SCO2 in the conduit 500. Two transducers504 and 506, such as piezoelectric transducers, are mounted on theconduit 500 opposing each other. Transducer 504 is a transmittingtransducer and transducer 506 is a receiving transducer. According tothe current invention, transducer 504 generates a sonic pulse 508 thatis transmitted through the stream 502 of SCO2 and is received bytransducer 506. By measuring the time lapse between the transmission ofthe sonic pulse 508 by transducer 504 and the reception of the pulse bytransducer 506, and taking into account the distance between thetransducers, the speed of sound of the SCO2 can be determined. Inaddition, an adjustment can be made to account for the velocity of theSCO2 through the conduit 500 by considering that the distance traversedby the sound wave is equal to the sum of the squares of the pipediameter and the distance down the pipe the flow has to travel duringthe signal interval. In particular, the flow velocity can be determinedby measuring the flow rate of the SCO2, for example using the flow meter32 shown in FIG. 1, and dividing the measured flow rate by the insidediameter of the conduit. By cross referencing the measured pressure ofthe SCO2 and the calculated speed of sound with data for the speed ofsound versus temperature at the measured pressure, such as that shown inFIG. 17, the temperature of the SCO2 can be accurately determined.

Regardless of the method used, preferably, the temperature of the SCO2is measured within ½ m of the inlet of the compressor 20.

Although the temperature measuring methods have been described above inconnection with a fossil fuel fired, dual cycle, supercritical fluid-airsystem for generating shaft power, it should be understood that themethod is equally applicable to other supercritical fluid systems, suchas an SCO2 system used in conjunction with a nuclear or solar heatsource.

As previously discussed, a challenge to implementation of any SCO2 cyclearises because of the very high pressures required (e.g., over 7.0 MPa)in order to achieve a supercritical condition. Such high pressures inthe SCO2 turbine makes sealing of the shaft extending from the turbineto the driven load difficult. As previously discussed, one approach isto incorporate the driven load into the SCO2 turbine pressure vessel.For example, the electric generator 90 in the FIG. 6 embodiment could beincluded within the SCO2 power turbine pressure vessel so that there isno need for extending a difficult to seal rotating shaft through thepressure vessel wall. However, as discussed above, such approach is notapplicable to situations in which the driven load cannot be includedwithin the SCO2 pressure vessel, such as a turboprop. Further, even whenthe approach is applicable, as in the case of electric generator, it hasdrawbacks.

According to one aspect of the current invention, a means is providedfor transmitting shaft power across the SCO2 turbine pressure vesselboundary without the need for sealing a shaft that penetrates thepressure vessel. As shown in FIGS. 19-23, an eddy current torquecoupling, or induction coupling, 36 is used to transmit power from theshaft 17, which is driven by the SCO2 power turbine 12″, to the shaft 58that, for example, drives an electric generator or turboprop. The shaft17, which is the input shaft for the torque coupling 36, rotates withinthe SCO2 power turbine housing 618 supported on bearings 622. Aninduction rotor 614 is affixed to, and rotates with, the shaft 17. Theinduction rotor 614 is made from a magnetically permeable material, suchas copper or aluminum.

A pressure membrane 612 attached to the housing 618 seals the SCO2within the housing. In a preferred embodiment of the invention, thepressure membrane 612 has a spherical curvature with the high pressureof the SCO2 in the housing 618 existing on the outside of the sphericalsurface. This places the membrane 612 in compression, which allows forthe use of materials that have substantially greater compressive thantensile strength, thereby allowing the membrane to made relatively thin.The thinness of the membrane 612 minimizes the gap between the armatures624, 626 and the induction rotor 614, which allows for greater torquetransmission. In a particular preferred embodiment of the invention, thepressure membrane 612 is made from a ceramic material such as, forexample, silicon nitride, which has excellent compression strength.

The housing 618 has an inlet port 602 in flow communication with aninlet manifold 604 and an outlet port 610 in flow communication with anoutlet manifold 608. Passages 606 connect the inlet and outlet manifolds604 and 608.

The shaft 58, which is the output shaft of the torque coupling 36,rotates within an armature housing 616 supported by bearings 630 and632. An armature assembly is coupled to the shaft 58. The armatureassembly comprises a bolt 640 that supports a first armature 624 withsouth facing magnetic poles and a second armature 626 with north facingmagnetic poles that are interleaved with the south facing poles of thefirst armature. The first and second armatures 624 and 626 arepreferably made from any appropriate paramagnetic material, such as, forexample, supermalloy. A permanent magnet 628, such as a neodymiummagnet, is supported on the bolt 640 radially inboard of the armatures624 and 626. The magnet 628 creates magnetic flux that extends betweenthe alternating poles of the armatures 624 and 626.

Relative rotation between the permanent magnet 628, coupled to theoutput shaft 58, and the magnetically permeable material of theinduction rotor 614, which is coupled to the input shaft 17, causes arate of change of magnetic flux resulting in an eddy flow of current inthe induction rotor. This current produces an opposing magnetic fluxwhich opposes the change in magnetic flux and thereby serves to transmittorque across the pressure membrane 612 to the armatures 624 and 626.However, there is slippage between the two shafts such that the outputshaft 58 rotates more slowly than the input shaft 17. The torquetransmitted across the pressure membrane from the input shaft 17 to theoutput shaft 56 reaches a peak at a rotor speed difference of about80-100 RPM.

Note that, alternatively, coils could be used instead of the inductorrotor solid material, in which case the stator and rotor would bothrotate. The losses associated with slip could then be captured aselectric current. This approach would require a brush system to transmitcurrent to a non-rotating structure. In addition, by using coils andvarying the resistance in the coil circuit, the torque transmitted couldbe varied, which could be useful for dynamic control.

The eddy current generated in the induction rotor 614 creates heat. Aspreviously discussed in connection with the embodiment illustrated inFIG. 3, in one embodiment of the current invention, the valve 51 directsa portion 52 of the SCO2 compressor discharge 48 to the eddy currentcoupling 36 for cooling purposes. In particular, as shown in FIGS.19-23, the stream 52 of cooling SCO2 is directed through the inlet port602 in the housing 618 and flows through an annular manifold 604. Fromthe manifold 604, the stream 52 of SCO2 flows through a series ofpassages 606 spaced circumferentially around the housing 618 thatconnect the inlet manifold 604 to the outlet manifold 608. A series ofvanes 650 are distributed around the passages 606 to aid in the transferof heat from the induction rotor 614 to the stream 52 of SCO2. As thestream 52 of cooling SCO2 flows through the passages 606 it absorbsheat, thereby cooling the induction rotor 614. After exiting the outletmanifold 608, the now heated stream 54 of cooling SCO2 exits the housing618 via the outlet port 610. As shown in FIG. 3, the SCO2 stream 54discharged from the eddy current torque coupling 36 flows through heatexchanger 10′ in which it transfers heat to the SCO2 56 that will beexpanded in the SCO2 power turbine 12′.

The valve 51 that controls amount of cooling SCO2 that is delivered tocool the eddy current coupling 36, shown in FIG. 3, may be controlled inreal time using temperature probes measuring the temperature of thestream 53 of SCO2 exiting the heat exchanger 10″ and temperature of thestream 54 of heated cooling SCO2 exiting the eddy current coupling 36for control feedback. The objective is to create two streams ofappropriate temperature such that when they are combined they have theproper “mixed” temperature to enable proper operation of the heatexchanger 10′.

Thus, according to one embodiment of the current invention, the heatgenerated by eddy current that must be removed from the eddy currentcoupling 36 is not lost from the system but is used to pre-heat aportion of the compressor discharge SCO2 that will be expanded in theSCO2 power turbine 12′. Although the power turbine 12′ must be sized toaccount for the power loss in the eddy current coupling 36, such powerloss results in the generation of heat that is fully recovered by thesystem.

Although the torque transmission method has been described above inconnection with a fossil fuel fired, dual cycle, supercritical fluid-airsystem for generating shaft power, it should be understood that themethod is equally applicable to other supercritical fluid systems, suchas an SCO2 system used in conjunction with a nuclear or solar heatsource.

Thus, although the current invention has been illustrated by referenceto certain specific embodiments, those skilled in the art, armed withthe foregoing disclosure, will appreciate that many variations could beemployed. Therefore, it should be appreciated that the current inventionmay be embodied in other specific forms without departing from thespirit or essential attributes thereof and, accordingly, referenceshould be made to the appended claims, rather than to the foregoingspecification, as indicating the scope of the invention.

What is claimed:
 1. A method of generating shaft power in a systemcomprising an air cycle and supercritical fluid cycle, comprising thesteps of: a) burning a fossil fuel in air so as to produce a combustiongas; b) expanding said combustion gas in at least a first turbine so asto produce an expanded combustion gas, said expansion of said combustiongas generating shaft power; c) compressing a supercritical fluid in afirst compressor; d) flowing at least a portion of said compressedsupercritical fluid through a first cross cycle heat exchanger, andflowing said combustion gas through said first cross cycle heatexchanger so as to transfer heat from said combustion gas to saidcompressed supercritical fluid so as to produce a heated compressedsupercritical fluid; e) expanding at least a portion of said heatedcompressed supercritical fluid in a second turbine so as to produce anexpanded supercritical fluid, said expansion of said supercritical fluidgenerating additional shaft power; and f) flowing at least a portion ofsaid expanded supercritical fluid through a second cross cycle heatexchanger, and flowing said air through said second cross cycle heatexchanger prior to burning said fossil fuel in said air so as totransfer heat from said expanded supercritical fluid to said air.